1. Field of the Invention
This invention relates generally to hydrodynamic rotary seals, such as rotary shaft seals, for establishing a seal between a relatively rotating surface and a structure, which wedge a lubricant film between the seal and the relatively rotating surface to inhibit seal wear and to exclude contaminants from the dynamic sealing interface. More specifically the present invention concerns ring-like rotary seals of the interference type which are composed of resilient sealing material which are contained within seal grooves and provide a seal geometry that cooperates with the seal groove to provide resistance to becoming skewed or twisted within the seal groove.
2. Description of the Prior Art
The prior art hydrodynamically lubricated compression-type rotary shaft seals disclosed in U.S. Pat. Nos. 4,610,319, 5,230,520, 5,678,829, 5,738,358, 5,873,576 and 6,036,192 are known in the industry by the registered trademark "Kalsi Seals", and pertain to products of Kalsi Engineering, Inc. of Sugar Land, Tex. The prior art teaches that the installed width of the body of such seals is difficult to predict, and can vary considerably due to a number of factors, including tolerances, compression variations, and thermal expansion. The prior art also consistently teaches that the width of the seal groove that is provided for such seals must be larger than the worst case installed width of the seal body in order to prevent increased interfacial contact pressure at the dynamic sealing interface resulting from simultaneous radial and axial constraint, and in order to prevent impairment of the hydrodynamic film and associated seal wear from such increased contact pressure. The contact pressure at the seal to shaft interface is one of the most important factors relating to hydrodynamic performance of the seal because it influences film thickness.
In U.S. Pat. No. 4,610,319 FIGS. 1 and 1A, the installed width of hydrodynamic seal member 50 is illustrated as being narrower than the width of seal groove 52. In FIG. 1B of the same patent, the installed width of hydrodynamic seal member 61 is illustrated as being less than the width of circular seal groove 63.
Likewise, in all of the figures in commonly assigned U.S. Pat. Nos. 5,230,520, 5,678,829 and 5,738,358 which illustrate an installed hydrodynamic seal, the width of the seal body is less than the width of the seal groove; see U.S. Pat. No. 5,230,520 FIGS. 1, 2, 4, 6, 7, 8, 9, 10 and 12, U.S. Pat. No. 5,678,829 FIGS. 1A, 3A and 4A, and U.S. Pat. No. 5,738,358 FIGS. 1, 2A, 4 and 5A. Prior art teaching pertaining to the importance of having a seal groove width that is wider than the installed width of the body of a hydrodynamic seal is also discussed in considerable detail in commonly assigned U.S. Pat. Nos. 5,873,576 and 6,036,192.
Hydrodynamic seal sales and implementation literature has also consistently taught the importance of providing a seal groove width that is wider than the installed width of the seal body in order to prevent seal damage associated with simultaneous radial and axial seal constraint. For example, this subject has been discussed in the "Gland Width Considerations" portion of the Kalsi Seals Rotary Shaft Seal Catalog PN 362-1, beginning with the catalog issue of Dec. 1, 1993, which states "The axial width of a Kalsi Seals gland has to be designed to accommodate the width of the compressed seal. Four primary factors affect the compressed seal width; (1) seal material displaced axially as a result of radial compression, (2) seal material displaced axially by the thermal expansion of the elastomer, (3) volumetric swelling of the elastomer due to media exposure, and (4) seal tolerances. If the groove length is not large enough to accommodate the aforementioned factors, the interfacial contact pressure between the seal and the shaft can increase dramatically, and result in a drastic reduction in hydrodynamic lubrication, and a corresponding decrease in seal performance."
The prior art also teaches that in the absence of differential pressure, the hydrodynamic rotary shaft seals of the type disclosed in U.S. Pat. Nos. 4,610,319, 5,230,520, 5,678,829 and 5,738,358 may be subject to skew-induced wear from impingement of environmental abrasives. For example, the "Gland Width Considerations" portion of the Kalsi Seals Rotary Shaft Seal Catalog PN 362-1 states "Seals used in applications having no differential pressure may tend to "snake" in the gland due to the effects of circumferential compression and thermal expansion. If snaking is present during rotation, the sharp exclusionary edge on the environmental side of the seal sweeps the shaft and causes environmental media impingement upon the environmental end of the seal. If the environmental media contains abrasive particulates, the impingement may cause abrasive wear of the environmental end of the seal. Some of the abrasives may also be swept into the dynamic sealing interface and cause interfacial seal and shaft wear."
The skew-induced impingement wear mechanism, and a solution that requires a washer and a mechanical spring to help to stabilize the seal against skew and to accommodate width variations of the seal body resulting from seal tolerances and thermal expansion, are described in SPE/IADC Paper No. 37627. This method prevents skew-induced impingement wear in the absence of differential pressure, but the seal can be subjected to pressure-responsive travel within the seal groove if the environment pressure exceeds the lubricant pressure and creates a hydraulic force across the area of the seal which exceeds the spring force. Unless the spring force is very consistent about the circumference of the seal, the environmental pressure may cock the seal within its groove, causing skew-induced impingement wear.
Commonly assigned U.S. Pat. Nos. 5,873,576 and 6,036,192 describe the skew-induced impingement wear mechanism in detail, and describe the use of resilient spring projections which are integral with, and projecting from, the seal body. These resilient projections are intended to stabilize the seal against skew-induced impingement wear while accommodating changes in the width of the circular seal body resulting from seal tolerances, thermal expansion, and seal material displaced by varying seal compression.
Testing has shown that the seal geometry disclosed in U.S. Pat. Nos. 5,873,576 and 6,036,192 successfully prevents skew induced wear in the absence of pressure, as was intended, and as such represents an improvement over older seal designs for certain applications. However, if the environmental pressure exceeds the lubricant pressure, the incomplete support provided by the resilient spring projections can in some embodiments permit the differential pressure to deform the seal body within the seal groove such that the seal attains a twisted and/or locally skewed position which is less favorable to environmental exclusion.
Testing has also shown that certain embodiments of seals constructed per the teachings of U.S. Pat. Nos. 5,873,576 and 6,036,192 are subject to pressure-responsive travel in the seal groove if the environment pressure exceeds the lubricant pressure and creates a hydraulic force which exceeds the spring force of the resilient spring projections. When the environment pressure is then removed, the difference in friction between the static sealing interface and the dynamic sealing interface can result in momentary twisting of the seal, which can be conducive to environmental ingestion.
In the seals disclosed in U.S. Pat. Nos. 5,873,576 and 6,036,192, communication passages are provided that communicate past the resilient spring projections to the cavity formed by the lubricant side groove wall, the peripheral groove wall, and the resilient spring projections themselves. The communication passages typically takes the form of the circumferential spacing of the spring projections. Testing has shown that the communication passages must be kept small for best seal constraint. However, in high runout applications, a small passage is less than optimum for use with non-Newtonian lubricants such as grease because the viscous resistance of the lubricant is not well suited to the rapid flow required through the passages in response to rapidly occurring volumetric changes caused by runout. As operating temperatures increase, the passages unfortunately become even smaller and less suitable for non-Newtonian lubricants due to thermal expansion of the sealing material.
With liquids containing particulates are forced into a small cavity, and then expelled out, the liquid fraction is more easily expelled than the particulates, and the particulates tend to build up and become entrapped in the cavity, where they typically create a tightly packed mass. The seals of U.S. Pat. Nos. 5,873,576 and 6,036,192 are not particularly suitable for use with greases containing solid lubricant particles because the particles will tend to pack-up in the small communication passages, and in the cavity formed by the lubricant side groove wall, the peripheral groove wall, and the resilient spring projections. Such greases are commonly used in the oil well drilling industry to lubricate heavily loaded critical service bearings.
Relatively soft sealing materials are often desirable for low pressure hydrodynamic seals because they help to minimize interfacial contact pressure, to maximize hydrodynamic lubricant film thickness, and to minimize seal-generated heat. The differential pressure-induced twisting, skewing and seal travel phenomena discussed above in conjunction with the seals disclosed in U.S. Pat. Nos. 5,873,576 and 6,036,192 are unfortunately more severe when a relatively soft sealing material, such as an 80 durometer Shore A elastomer, is employed.